Hydraulic mounts are dual aspect devices. In a first aspect, an hydraulic mount provides location of one object, such as a motor vehicle engine, with respect to a second object, as for example the frame of the motor vehicle. In a second aspect, the hydraulic mount provides damping of vibration or low dynamic stiffness as between the first and second objects, as for example damping or isolating of engine vibration with respect to the frame of the motor vehicle.
Hydraulic mounts which are used for motor vehicle applications are represented, for example, by U.S. Pat. Nos. 4,828,234, 5,215,293 and 7,025,341.
U.S. Pat. No. 5,215,293 discloses an hydraulic mount having a rigid upper member which is bolted to the engine and a lower engine member which is bolted to the frame, wherein the upper and lower members are resiliently interconnected. The upper member is connected to a resilient main rubber element. Vibration of the main rubber element in response to engine vibration is transmitted to an adjoining upper fluid chamber. The upper fluid chamber adjoins a rigid top plate having an idle inertia track therethrough which communicates with an idle fluid chamber. The idle fluid chamber is separated from an idle air chamber by an idle diaphragm. The idle air chamber is selectively connected to atmosphere or to engine vacuum in order to selectively evacuate the idle air chamber in which case the idle diaphragm is immobilized. A bounce inertia track is formed in the top plate and communicates with a lower fluid chamber which is fluid filled. A bellows separates the lower fluid chamber from a lower air chamber which is vented to the atmosphere.
The idle inertia track has a larger cross-sectional area and a shorter length than that of the bounce inertia track, such that the ratio provides resonant frequency damping at the respectively selected resonance frequencies. In this regard, the resonance frequency of the fluid flowing through the idle inertia track is set to be higher than that of the fluid flowing through the bounce inertia track. As such, this prior art hydraulic mount is able to effectively damp relatively low frequency vibrations over a lower frequency range, such as engine shake or bounce, based on resonance of a mass of the fluid in the bounce inertia track, while, on the other hand, the idle inertia track is tuned so that the hydraulic mount exhibits a sufficiently reduced dynamic stiffness with respect to relatively high-frequency vibrations over a higher frequency range, such as engine idling vibrations, based on the resonance of a mass of the fluid in the idle inertia track.
In operation, vibrations in the higher frequency range are isolated by operation of the induced fluid oscillations in the upper fluid chamber passing through the idle inertia track and the resilient deformation of the main resilient element and the idle diaphragm in that the idle air chamber is at atmospheric pressure. For vibrations in the lower frequency range, the idle air chamber is evacuated by being connected to engine vacuum, wherein now the fluid oscillations of the upper fluid chamber travel through the bounce inertia track and are damped thereby in combination with the resilient deformation of the main resilient element and the bellows.
While hydraulic mounts work well, there is the problem that each of the tracks, the idle inertia track and the bounce inertia track, provide low dynamic stiffness and damping at a respective predetermined range of frequency of vibration. In particular with respect to the idle inertia track, the “idle rate dip” frequency is singular, being selected to generally suit a particular engine. For example as shown at FIG. 4B, which is a graph 300′ of dynamic stiffness of a prior art hydraulic mount versus engine vibration frequency (discussed further hereinbelow), wherein the plot 302′ has a singular idle rate dip 304′ occurring at a frequency of 50 Hz.
Engine idle vibration has more than one order (vibrations per revolution). For example, it is well known physics that the undamped natural frequency, fn, for a simple mass-spring system is given by:
                              f          n                =                              1                          2              ⁢                                                          ⁢              π                                ⁢                                                    k                /                m                                      .                                              (        1        )            where k is the spring stiffness and m the sprung mass.
Aspects of internal combustion engines which are relevant to the design of hydraulic mounts therefor include:                1. the dynamic stiffness K* (complex dynamic stiffness);        2. the elastic modulus (elastic dynamic stiffness) K′;        3. the out-of-phase modulus (loss modulus, or loss dynamic stiffness) K″, where term K*=√(K″2+K′2);        4. the loss angle θ=Tan−1(K″/K′);        5. the damping coefficient (C), where tan θ=Cω/K′, where ω=2πf, and where f=frequency;        6. the resonance frequency of the fluid column in the inertia track which is the frequency at which the out-of-phase module (K″) reaches a maximum;        7. the rate (dynamic stiffness) dip frequency, which is the frequency at which the K* reaches a minimum, wherein the rate dip frequency is several Hz (frequency) lower than the maximum K″ frequency;        8. the are two zones for noise and vibration (N and V): control and isolation;        9. the control zone, which is the frequency below 1.414×fn (natural frequency), wherein in the control zone, damping is required to reduce the vibration, engine bounce and shake or rough road shake (e.g., at low frequency—control zone) requires high damping to reduce the vibration;        10. the isolation zone, which is the frequency above 1.414×fn, wherein in the isolation zone, low dynamic stiffness is required to isolate the vibration, engine idle shake (e.g., at high frequency—isolation zone) requires low dynamic stiffness to isolate the vibration, wherein the isolation zone, the damping increases the vibration;        11. the natural frequency of the fluid in the inertia track, which depends on the mass of the fluid in the track and the stiffness (bulge/volume stiffness of the upper chamber and lower chamber); and        12. the engine RPM at idle, for example at 900 RPM, the first order frequency is 15 Hz (e.g., 900/60=15), and wherein the engine firing orders/frequencies of different engines are shown in the following Table I.        
TABLE IEngine1x firing2x firing3x firing . . .I42nd order (30 Hz)4th order6th order. . .V63rd order6th order9th order . . .V84th order8th order12th order . . .
Examples of the one times and two times engine firing frequencies are: for a V6 internal combustion engine, the 3rd and 6th engine vibration orders have the most undesirable vibration, and for a V8 engine, the 4th and 8th engine vibration orders have the most undesirable vibration. The selection of the tuned idle rate (e.g., dynamic stiffness) dip frequency is, therefore, a best compromise selection of the frequency of the idle rate dip the one times and two times engine firing frequencies, as shown at FIG. 4B.
The vacuum tubing interconnecting with the idle air chamber introduces an undesirable noise factor in the sense of an unwanted dynamic stiffness. Therefore in the prior art, the existence of the vacuum tubing is considered problematic, and there is a general need to introduce countermeasures to reduce the dynamic stiffness caused by the vacuum tubing, or move the unwanted dynamic stiffness to a frequency range which is not under consideration with respect to optimization of the idle rate dip.
Accordingly, it would be desirable in the art if somehow a double idle rate dip frequencies of dynamic stiffness could be provided, and further, if somehow the unwanted dynamic stiffness of the vacuum tubing could somehow be eliminated as a problem.